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United States Patent |
5,575,737
|
Weiss
|
November 19, 1996
|
Utility vehicle driveline control method
Abstract
A method is described for controlling the power train of a work vehicle.
The power train includes an engine with a fuel injection control and an
infinitely variable transmission. The method include first determining an
initial engine speed, an initial wheel speed, the actual engine speed, and
the actual wheel speed. To improve comfort and reduce fuel consumption and
emissions two steps are performed. First, the transmission ratio of the
infinitely variable transmission will be increased if the actual engine
speed is less than the initial engine speed, or decreased if the actual
engine speed is greater than the initial engine speed. Second, while the
transmission ratio remains constant, the engine speed will be adjusted
until the engine speed is the same as the initial engine speed.
Inventors:
|
Weiss; Heinz (Bensheim, DE)
|
Assignee:
|
Deere & Company (Moline, IL)
|
Appl. No.:
|
517482 |
Filed:
|
August 21, 1995 |
Foreign Application Priority Data
| Aug 27, 1994[DE] | 44 30 447.1 |
Current U.S. Class: |
477/43; 477/44; 477/46 |
Intern'l Class: |
B60K 041/12 |
Field of Search: |
477/37,43,44,46,107,108,110
|
References Cited
U.S. Patent Documents
4039061 | Aug., 1977 | Pruvot et al. | 477/37.
|
4622867 | Nov., 1986 | Nishioka et al. | 477/46.
|
4648040 | Mar., 1987 | Cornell et al.
| |
4653005 | Mar., 1987 | Osanai et al. | 477/43.
|
4671138 | Jun., 1987 | Nobumoto et al. | 477/46.
|
4735112 | Apr., 1988 | Osanai et al. | 477/43.
|
4872115 | Oct., 1989 | Itoh et al. | 477/43.
|
5009127 | Apr., 1991 | Morimoto et al. | 477/46.
|
5009129 | Apr., 1991 | Morimoto et al. | 477/46.
|
Foreign Patent Documents |
0280757 | Sep., 1988 | EP.
| |
0415048 | Mar., 1991 | EP.
| |
3508155 | Sep., 1985 | DE.
| |
3628490 | Apr., 1987 | DE.
| |
3533193 | May., 1987 | DE.
| |
4115623 | Nov., 1992 | DE.
| |
4223967 | Jan., 1994 | DE.
| |
Primary Examiner: Marmor; Charles A.
Assistant Examiner: Kwon; Peter
Claims
I claim:
1. A driveline control system for a utility vehicle having an engine (10)
with fuel injection control system, a manually-operated throttle lever,
and an infinitely variable transmission (16), the driveline control system
having means for determining and storing in memory an initial engine
output rotational speed (N(mo)), an initial wheel rotational speed
(N(ro)), an actual engine rotational speed (N(mn)) and an actual wheel
rotational speed (N(rn)), the driveline control system performing a method
comprising the following steps:
increasing the transmission ratio of the infinitely variable transmission
if the actual engine speed N(mn) is less than the initial engine speed
N(mo);
decreasing the transmission ratio of the infinitely variable transmission
if the actual engine speed N(mn) is greater than the initial engine speed
N(mo); and
while the transmission ratio remains constant, adjusting the engine speed
until the actual wheel rotational speed N(rn) is the same as the initial
wheel rotational speed N(ro).
2. The method of claim 1, wherein:
the steps are repeated at regular time intervals.
3. The method of claim 1, wherein:
the initial wheel rotational speed (N(ro)) is replaced by a target wheel
rotational speed value (N(rs)) which can be pre-set by an operator.
4. The method of claim 1, wherein:
in response to a change in engine rotational speed which results from a
change to a new load, and by means of a stored engine characteristic in
the form of lines d of constant fuel injection pump position, the control
system determines a new load hyperbola of constant load corresponding to
the new load and determines an intersection of the new load hyperbola with
a pre-set increased speed control line or a pre-set full-load control
line; and
the control system adjusts the drive ratio of the transmission and adjusts
the engine rotational speed in order to operate at favorable operating
points of the engine performance map with respect to fuel consumption and
emissions.
5. The method of claim 4, wherein:
the favorable operating points lie on an increased speed control line that
can be pre-set or on a full-load control line that can be pre-set.
6. The method of claim 1, further comprising:
pre-setting and maintaining a desired engine rotational speed when the
engine is powering an auxiliary devices.
7. The method of claim 1, characterized by:
in response to adjustment of the throttle lever, reducing an adjustment
range of an automatic efficiency improvement so that defined engine
rotational speeds are modified to accomodate an auxiliary drive unit
driven by the engine.
8. The method of claim 1, wherein:
an auxiliary device is driven by the engine; and
adjusting the transmission drive ratio in response to a changing load on
the engine resulting from the auxiliary device to maintain a constant
engine rotational speed while modifying the wheel rotational speed.
9. The method of claim 1, wherein:
a) the transmission drive ratio is raised or lowered according to the
relation I(n)=I(o).times.N(mo)/N(mn), wherein I(o) is the original drive
ratio and I(n) is a new transmission drive ratio; and
b) the engine rotational speed is adjusted according to the equation
N(ms)=I(n).times.N(ro).times.N(rs)/N(ro), wherein which N(rs) is taken as
N(ro) the initial wheel rotational speed and N(ms) is a new target
rotational speed of the engine.
Description
BACKGROUND OF THE INVENTION
The invention concerns a process and a control system for the control of
the driveline of a utility vehicle that contains a power plant with fuel
injection quantity control and an infinitely variable transmission (IVT),
in which an initial engine rotational speed, an initial wheel rotational
speed, the actual engine rotational speed and the actual wheel rotational
speed are determined and stored in memory.
The driveline of utility vehicles, such as agricultural or commercial
vehicles, generally contain an internal combustion engine and a
transmission. The engine has the requirement that it supply the needed
power with the best possible efficiency, while the transmission operates
as torque and rotational speed converter adjusting the performance map of
the engine to the demand map of the vehicle. In the further development of
the driveline, beyond the aforementioned goals of increased productivity,
operator comfort, other aspects come increasingly into the foreground,
such as exhaust emissions and fuel consumption.
In order to attain these goals infinitely variable transmissions may be
employed to great advantage, such as have been described in DE-A-35 33 193
and DE-A-41 15 623. These are hydrostatic-mechanical torque dividing
transmissions with an infinitely variable hydrostatic component,
consisting of adjustment pump and hydraulic motor, and a mechanical branch
with several drive ratios that can be shifted automatically without
interrupting the power flow. The drive ratio of these transmissions is
infinitely variable over the entire operating range.
EP-A-0 280 757 describes a control and regulating arrangement for such an
infinitely variable transmission. The actual engine rotational speed and
the actual transmission output rotational speed are continuously
determined and compared to the target signal. By continuously varying the
drive ratio of the transmission and the engine rotational speed the
control system reacts to changes in the target signals or the operating
conditions. With increasing tractive resistance the engine rotational
speed is initially reduced which results in an increase in the engine
control signal in the drive control, in order to make available an
increase in engine power. Furthermore the transmission drive ratio is
adjusted in order to maintain or to attain the desired vehicle speed. The
control system is to be designed in such a way that the transmission drive
ratio as well as the engine rotational speed are continuously adjusted for
optimum fuel economy.
Furthermore in a technical meeting in Dresden, Germany, in 1989 F. Jarchow
proposed an infinitely variable hydrostatic-mechanical transmission for
tractors that can be shifted under load and has a control that permits
operation along a curve of minimum fuel consumption. On the basis of an
optimum fuel consumption curve a voltage can be determined for each
position of the gas pedal, which is compared with a voltage corresponding
to the transmission input rotational speed. The voltage difference is used
to adjust the transmission drive ratio.
SUMMARY OF THE INVENTION
Accordingly, an object of this invention is to provide an improved control
system for the control of the driveline of utility vehicles of the
aforementioned type.
A further object of the invention is to provide such a control system
through which the engine can be operated with regard to main drives.
Another object of the invention is to provide such a control system through
which the engine can be operated with regard to secondary drives for
hydrostatic pumps, mechanical power take-off shaft drives and the like in
terms of productivity, fuel economy and emission characteristics without
additional effort by the operator in a favorable operating regime.
These and other objects are achieved by the present invention, which
permits the attainment of the goal of reduced fuel consumption with
simultaneously reduced emissions as well as an increase in operator
comfort. The suggested strategy of transmission drive ratio adjustment
permits operation of the engine in areas of the performance map with
favorable efficiency. Beyond that it attains improved productivity
combined with higher economic efficiency.
In order to utilize fully the potential of the engine, the engine is
operated along the constant output torque hyperbola by increasing
(decreasing) the transmission drive ratio with constant vehicle
parameters, such as speed and load, where the load increases and the
engine rotational speed decreases.
It was recognized in particular that the tractive force hyperbola of the
vehicle can be normalized and reproduced by means of the constant axle
reduction gear ratio on the hyperbolas of constant power of the driveline,
so that the areas and the relations of the rectangles to the points of the
hyperbolas can be easily estimated and compared to each other.
For reasons of cost, the initial concept is based on a mechanical fuel
injection pump which is controlled primarily by the electronic control of
the transmission with respect to the determination of the quantity of fuel
injected. The proposed process, however, can be transferred without any
problems to electronically controlled fuel injection pumps.
Since the operating range of the engine is described only in terms of
parameters specific to the engine, the process for the control of a
driveline can be applied to all infinitely variable transmissions with
torque division and covers the purely mechanical infinitely variable
torque dividing transmission Torotrack of the firm Leyland as well as all
purely hydrostatic or hydrostatic-mechanical transmissions with torque
division. This permits a simple and flexible control of transmission and
engine, without requiring torque sensors.
BRIEF DESCRIPTION OF THE DRAWINGS
FIG. 1 is a block diagram of a control system according to the invention.
FIG. 2 shows an example of the characteristic of an engine with a
definition of the curves of the control of the driveline.
FIGS. 3, 4 and 5 show further engine characteristics relating to the
invention.
DETAILED DESCRIPTION
FIG. 1 shows an engine 10 with variable rotational speed that can be
controlled by a controller 12 which preferably includes as microprocessor.
The controller 12 is preferably connected through a control area network
bus 14 or over a direct connection to a transmission controller 18, from
which it receives control signals for the fuel injection pump (not shown)
of the engine 10 and transmits them to the latter. Thereby the fuel
injection pump is not connected directly to the gas pedal or the
manually-operated throttle lever 21, as is usual in conventional vehicles,
but is controlled by an electronic regulating unit. This regulating unit
determines the amount of fuel injected considering the operator's desires
(lever and pedal positions), the operating conditions and the performance
map of the engine 10.
The engine 10 drives an infinitely variable transmission 16 without an
intervening clutch. The drive ratio of the transmission 16 is determined
by a drive ratio controller 18, also preferably including a microprocessor
and which is also connected to the bus 14. The electronic control of the
fuel injection pump 12 and the controller 18 of the transmission 16
interact with each other and may be combined into one component if
desired. Furthermore, the bus 14 is connected to an operator's panel 20
through which the operator can provide inputs to influence the control of
the driveline.
The infinitely variable part of the transmission 16 preferably consists of
an adjusting pump (not shown) with a hydraulic motor (not shown). In the
mechanical part several drive ratios are provided which can be shifted
automatically without interruption of the power flow. Through the
infinitely variable part the transmission permits starting from standstill
without the need for a starting clutch. For the standstill condition or
the braking of the vehicle, clutches (not shown) separate the driveline
from the driving wheels (not shown), in order to avoid stalling the engine
or to avoid working of the transmission against the brakes. After the
separating process is completed, the separating clutches engage again
automatically when synchronization is attained. In place of such a
transmission other infinitely variable transmissions may be applied.
Electronic rotational speed measurement transmitters detect the engine
output rotational speed N(m) and the transmission output rotational speed,
which corresponds to the normalized wheel rotational speed N(r) of the
drive wheels. The rotational speeds N(m) and N(r) are continuously
measured and stored in memory by the electronic control unit. From these
the operating points of the engine and transmission are determined
depending upon the pre-set target values.
A hydraulic pump 22 is coupled directly by gears to the output of the
engine 10 and supplies the vehicle components and the attached implements
as well as an implement power take-off shaft 24. Therefore these rotate in
proportion to the engine rotational speed.
In contrast to conventional drive concepts, the operator of the vehicle
described here with infinitely variable transmission only has the
possibility of providing the input for the target values for the velocity
and engine rotational speed etc. through an electronic control unit (drive
management). A direct influence on the engine operating point or the
transmission drive ratio is no longer possible.
The operator provides the input of the target value and the electronic
control determines the manner in which these targets are to be reached.
The input of the target value for the vehicle speed can be made through a
velocity selector lever, a gas pedal or a retarder circuit integrated into
the brake pedal, which are not shown here but have been described, for
example, in WO-94/06651. The vehicle speed input is provided with the
selector lever. The use of the retarder or the gas pedal permits a
modulation of the target velocity, where the foot-operated gas pedal
permits an increase in the velocity or maintains that velocity despite an
increase in the load, while the retarder decelerates the vehicle without
friction devices. An additional hand-operated throttle lever establishes a
particular engine rotational speed that either must be maintained as
constant (power take-off shaft operation) or that may not be underrun.
The performance map of an engine shown in FIG. 2 characterizes an engine
that is not yet optimized for an application with an infinitely variable
transmission. This can be seen in the fact that the point K of the optimal
specific fuel consumption is to the right of the point M(max) of the
maximum engine torque.
The engine torque M(m) was plotted against the engine output rotational
speed N(m) and the full-load torque curve M(voll) was entered. Furthermore
the torque hyperbolas of constant power output P(i) including the rated
power output PN are entered at the associated rated rotational speed N(N)
as well as the lines of specific fuel consumption b and lines d of
constant fuel injection pump position.
The lines d result from measurement points of the actual engine at
different constant fuel injection pump positions. The actual d lines may
be curved, but they are linearized and are stored in the memory of the
controller as the straight lines d.
According to FIG. 2 the operating area of the engine is defined by the
following lines or curves:
the torque hyperbola of constant power output (P4=40%) of the special
application case, which can be determined from full load rotational speed
and actual rotational speed of the engine by means of the stored lines d
of constant fuel injection pump position (proportional torque),
the increasing speed control line 1, which has been established to define
the smoke limit in the engine performance map is preferably a
straight-line torque curve between the points of the lower idle rotational
speed and the maximum torque M(max) of the full-load curve M(voll),
the full-load control line 2, which is also preferably established as a
straight line and defines the variation of the torque curve in the region
of the most favorable fuel consumption, and
a momentary performance map, according to which the engine, at a given
quantity of fuel injected, reacts to a change in load.
Infinitely variable transmissions without a main clutch have the possible
disadvantage that they cannot extract any additional energy from the
flywheel during the start-up process and are therefore considered poor to
react without special precautions. In order to counteract this behavior,
the IVT does not operate the engine along the increasing speed control
line 1 to control its acceleration or even to the left of this curve in
the smoke region, but operates intermediate engine speeds, if necessary by
means of the hand-operated throttle lever, in order to be prepared for
unknown load demands.
Corresponding, for example, to a the very heavy line P in FIG. 2, the idle
rotational speed was set at 2000 r.p.m. Here the mechanical fuel injection
pump is able, through its internal characteristics, to automatically meet
the demands of a load increase, whereby the rotational speed of the engine
was reduced from 2000 r.p.m. to a value of 1920 r.p.m. Since the lines d
of constant fuel injection pump position are stored in the memory of the
IVT controller, it is possible to determine in this way the proportion of
the load as x=40% of the rated torque.
The 40%-torque hyperbola is thereby defined and has a definite intersection
with the increased speed control line 1, so that the engine rotational
speed can be adjusted incrementally in the direction towards this point.
For this purpose several calculation cycles may be required, where
simultaneously the general level of the load may also change from cycle to
cycle. The mechanical fuel injection pump can follow these demands without
any problem, so that starting from the instantaneous lines d of constant
fuel injection pump position, a known value for the target rotational
speed and the measured value for the actual rotational speed, the input
parameters for the next calculation cycle are available, as was explained
in the preceding section.
If in the meantime, a minimum speed for the engine was given as input by
the manually-operated throttle lever, then the intersection of the
hyperbola is not calculated with the increasing speed control line 1, but
with the vertical S to the rotational speed, which was provided as input
by the manually-operated throttle lever. Here too, the process proceeds
incrementally. From the polygon enclosed by the speed increasing control
line 1, the full-load straight line 2; the lines d of constant fuel
injection pump position and the load hyperbola, the area of the possible
improvement in efficiency can be estimated (efficiency polygon). It can be
seen clearly that the manually-operated throttle lever adjustment has
almost halved the area of the efficiency polygon.
With power take-off shaft operation the same procedure is used with the
difference that the manually-operated throttle lever adjustment is shifted
further to the right into the vicinity of the line P or the design
rotational speed of the power take-off shaft, in order to provide an input
of a constant power take-off shaft rotational speed. Thereby the
efficiency polygon has been reduced almost to a point.
The efficiency polygon and the rectangle starting at point A under the
torque hyperbola thereby provide a good overview over the status of the
driveline with respect to power output and efficiency.
In an engine preferred for application with IVT the point for the most
favorable fuel consumption is at the extreme left and upward in the engine
performance map (that is, at low engine rotational speed), in order to
attain a great reduction in engine rotational speed and thereby gain a
large improvement in mechanical efficiency of engine and transmission and
to further improve efficiencies at optimum thermal loading of the engine.
At constant torque the full load straight line 2 lies preferably above the
point of the rated load PN and begins at rotational speeds less than N(N),
in order to make possible a closer accommodation of the performance map of
the engine with that of the transmission. The accommodation in the region
of the most favorable fuel consumption does not appear critical as long as
the shell-shaped curves represent ellipses lying flat. Most appropriately,
the full-load straight line lies 10% to 15% below the full load curve
M(Voll), in order to retain a torque reserve for peak loads during heavy
traction operations and, on the other hand, in order to avoid overloading
the transmission over a longer period of time.
For relatively small loads the full-load operating point of the engine is
located along the increased speed control line 1. With a further increase
in the torque the engine behaves as already described. With a constant
fuel injection quantity the engine increases its torque along the line P
and the operating point moves to the left beyond the increasing speed
control line 1.
Since the logic for the determination of the engine target and actual
values has not changed, the associated hyperbola for constant power can be
determined as described. When looking at the intersection of the hyperbola
with the increasing speed control line 1 it is evident, however, that the
hyperbola lies above the original hyperbola and can be brought to an
intersection with the increased speed control line 1 only by increasing
the rotational speed. The result is an intersection of the new hyperbola
with the increased speed control line 1 at a higher level. In order to
maintain the desired vehicle velocity an adjustment of the transmission
drive ratio is necessary.
According to FIG. 2 the process described meets the load demands up to
approximately 63% of the rated output. The hyperbolas greater than 63% no
longer intersect with the increased speed control line 1. Instead they
intersect with the full-load straight line 2, so that these intersections
must be distinguished from each other.
For power output values greater than 63%, the engine operating point lies
along the full-load straight line 2. An increase (decrease) in the load or
the speed therefore requires an increase (decrease) in the power output of
the engine, that is, the engine operating point is shifted along the
full-load straight line 2 further to the right (left). This, again, occurs
in incremental steps by an increase (decrease) in the engine rotational
speed. Below the full-load straight line 2, however, all possibilities for
improvement in efficiency are available by a change in the transmission
drive ratio and a reduction in speed with a simultaneous increase in the
torque in the direction toward the full-load straight line. The reduction
(increase) in resistance connected with the reduction in speed may
eventually require a further adjustment in the transmission drive ratio.
The speed reducing control line 3 for the rotational speed limitation of
the Diesel engine protects the engine against excessive rotational speed,
defects in the electronic control or against improper operation by the
operator (for example, wide-open throttle setting at idle speed). This is
attained by mechanical means in the mechanical fuel injection pump, but is
being increasingly attained electronically in future fuel injection pumps.
While an increase (decrease) in vehicle power starting from a low power
level (increased speed control line 1) requires a decrease (increase) of
the transmission drive ratio with simultaneous increase (decrease) in the
engine rotational speed, an increase in vehicle power starting from a high
power level (full-load line 2) signifies an increase (decrease) in the
transmission drive ratio with a simultaneous increase (decrease) in the
engine rotational speed.
By infinitely variable changes in the transmission drive ratio the IVT
permits operation over the entire speed range with constant engine output.
As revealed by FIG. 3, the engine operating point X can be varied by
adjusting different transmission drive ratios i in a region Y marked by
arrows of the traction force hyperbola which lies between a minimum wheel
rotational speed nrmin and a maximum wheel rotational speed nrmax. For
this reason a vehicle with IVT could do without a constant power
characteristic of the engine. Arguments for the retention of the constant
power characteristic include: Greater penetration force for heavy power
take-off shaft implements and a possible consolidation of the performance
maps of the engine and the transmission.
The power output may be represented as an area in the torque/r.p.m.
diagram. Due to the relationship "P=M.times.n" all areas F1, F2 are
located below a power hyperbola, are equal in area and represent a measure
of the power output (FIG. 3). This also applies to the areas of the
differential power output which, for example, must be made available by
the engine for an increase in the load (FIG. 4).
For the sake of simplicity the efficiency .eta. was initially disregarded
or set equal to 1. Limitations of this type, however, do not exist, since
the efficiency cancels out in the following equations.
Depending upon the magnitude of the disturbance and the control input, the
necessary adjustments for a new operating point (IVT drive ratio, engine
r.p.m.) are determined by appropriate control strategies. For this purpose
the engine rotational speed nm and the output rotational speed nr are
continuously measured and stored in memory. In order to recognize a change
in the load, it is necessary to permit a decrease (increase) in the wheel
speed or the engine speed, and to register it, before the control process
can be brought into action.
On the basis of FIG. 5 the control strategy during load increase (engine
depression) is explained as follows. The engine is operated along the
full-load line, that is, at M(mo)=constant. In its final effect the
desired vehicle velocity and therefore also the wheel rotational speed nro
is to be held constant. The temporary reduction in wheel rotational speed
due to engine depression is seen as negligible and is corrected in the
course of the control process.
The following abbreviations are used:
______________________________________
N(mo) engine output rotational speed (measured)
N(mn) new engine rotational speed (measured)
N(ms) engine target rotational speed (calculated)
N(ro) wheel output rotational speed (measured)
N(rn) new wheel rotational speed (measured)
N(rs) wheel target rotational speed (pre-set)
M(mo) constant engine output torque
M(ro) wheel output torque
M(rn) new wheel torque
I(o) IVT drive ratio at initial point (calculated)
I(n) new IVT drive ratio (calculated)
______________________________________
According to original assumption, the input and output power of the
transmission are connected to each other as follows:
.eta..times.2.pi..times.M(mo).times.N(mo)=2.pi..times.M(ro).times.N(ro)
(1).
Due to the equal areas of the rectangles of FIG. 5 the above relationship
also holds for the differential power output:
.eta..times.2.pi..times.(N(mn)-N(mo)).times.M(mo)=2.pi..times.N(ro).times.(
M(rn)-M(ro)) (2).
It follows therefrom for a shift from (I) to (III) for the transmission or
from (IV) to (V) for the engine:
N(mn)=N(mo).times.M(ro)/M(rn) (3)
and
I(n)=N(mn)/N(rn)=N(mo)/N(ro).times.M(ro)/M(rn) (4).
Since these equations still contain a torque relationship, but for
technical reasons only rotational speeds are being measured, it is
necessary to divide the process into two phases.
In the first phase reaction is made to the engine depression observed
(N(mn)<N(mo)), in that the transmission drive ratio (I=N(m)/N(r)) of the
IVT is increased and thereby the wheel rotational speed N(r) is reduced,
until the initial engine depression is compensated for and the engine
output rotational speed N(mo) is again reached. Due to the increase in the
transmission drive ratio the operating point of the wheel torque is
shifted along the torque hyperbola shown in FIG. 5 from (I) to (II) and
thereby from M(ro) to M(rn). This compensation insures that the engine
operating point (IV) remains unchanged.
In the second phase the engine rotational speed (by increasing the fuel
injection quantity) and the wheel rotational speed are raised, with
constant engine and vehicle torques, until the original vehicle velocity
is again reached, it had been lowered by the transmission drive ratio
adjustment of the first phase. Thereby the engine power output as well as
the wheel power output are increased and the operating points in FIG. 5
move from (IV) to (V) for the engine or from (II) to (III) for the
transmission.
The goal is to attain the least possible deviation from the target
rotational speed of the wheels (here: N(r)s=N(ro)). Therefore a process
control is performed after phase one and two in brief time intervals
and/or when the engine rotational speed is underrun or exceeded by a
pre-set amount as a result of a change in the vehicle load.
In the first phase, in particular, the procedure is as follows:
With a load increase (see the example illustrated in FIG. 5) N(mo)/N(mn)>1.
As long as N(mn)<N(mo) the IVT drive ratio is increased, whereby the wheel
rotational speed is decreased. The following relationship can be derived
for the new transmission drive ratio from the above equations:
I(n)=I(o).times.N(mo)/N(mn) (5).
If the load is unchanged, then N(mo)/N(mn)=1. Therefore it follows that
I(n)=I(o) and no change is performed on the drive ratio of the IVT.
When the load is reduced, then N(mo)/N(mn)<1. As long as N(mn)>N(mo) the
IVT drive ratio is reduced whereby the wheel rotational speed is
increased. The new drive ratio is here also calculated from the equation
(5).
In the second phase, in particular, the procedure is as follows:
If the desired speed or load are increased (N(rs)/N(ro)>1) by means of the
adjustment of the IVT in the first phase, then the wheel rotational speed
N(rn) must change to comply. This is performed by an adjustment in the
engine rotational speed at constant I(n) (as long as the speed decreasing
control line 3 had not been reached). The change in the engine rotational
speed is performed according to the following relation which can be
derived from the above equations:
N(ms)=I(n).times.N(ro).times.N(rs)/N(ro) (6).
If the speed and load remain unchanged, then N(rs)=N(ro) and thereby
N(ms)=N(mo), and there is no change in the engine rotational speed.
When the load is reduced N(rs)/N(ro) will be less than 1. The engine
rotational speed is reduced according to the relation (6) as long as the
increased speed control line 1 on the left side had not previously been
reached.
While the present invention has been described in conjunction with a
specific embodiment, it is understood that many alternatives,
modifications and variations will be apparent to those skilled in the art
in light of the foregoing description. Accordingly, this invention is
intended to embrace all such alternatives, modifications and variations
which fall within the spirit and scope of the appended claims.
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